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INVESTIGATION OF POWERTRAIN MODES

Basem Alzahabi Arnaldo Mazzei Logesh Kumar Natarajan Associate Professor Assistant Professor Graduate Research Assistant Dept. Mechanical Engineering Dept. Mechanical Engineering Dept. Mechanical Engineering Kettering University Kettering University Kettering University 1700 West Third Avenue 1700 West Third Avenue 1700 West Third Avenue Flint, Michigan, 48504, USA Flint, Michigan, 48504, USA Flint, Michigan, 48504, USA

Abstract Nomenclature

Powertrain mounting systems serve a number of functions J of inertia of powertrain with respect for the overall vehicle. First and foremost, the mounting xx to X-axis of powertain inertia co-ordinate system. system maintains the of the powertrain in the vehicle as it is subjected to inertia and loads. In J yy mass of powertrain with respect addition, the mounting system controls the overall motion of to Y-axis of powertain inertia co-ordinate system. the powertrain, preventing the engine, transmission, and mass moment of inertia of powertrain with respect accessories from contacting other components of the J zz vehicle, thereby avoiding damage to any vehicle systems to Z-axis of powertain inertia co-ordinate system. from the potential impact. Introduction The mounting system also plays a role in vehicle ride and handling due to the significant mass of the powertrain that Powertrain mounting systems are required to maintain the can be approximately 15 to 20% of the total mass of a large position of the vehicle’s powertrain with respect to its body car. Finally, the mounting system isolates the rest of the and components as the vehicle is subject to several loads. vehicle and its occupants from the powertrain as a vibration These loads (to name a few) are due to road input, vehicle source. Therefore, investigating the powertrain rigid body , braking and cornering. Within tolerances, the modes for a particular mounting system is very critical in position of the powertrain is controlled by its mounting determining initial mount locations and rates for a new system, so the powertrain does not interfere with the powertrain or vehicle architecture. vehicle’s structure. Also, the mounting system is responsible for isolating engine vibration from the vehicle and from its This paper investigates the design of a powertrain mounting passengers. The amount of isolation and NVH performance system of a front wheel drive vehicle with a transversely of the mounting system will depend on the vehicle design mounted V8 engine and mounted on an isolated constraints. For instance, if the engine compartment does front sub-frame. A torque strut is attached to the top of the not allow for enough room for powertrain motion, vibration engine and to the body structure with elastomeric bushings isolation may not be best due to high stiffnesses required for with torque reaction as its function. The locations of the mounts in order to prevent contact between engine and the powertrain mounts and torque strut cannot be modified vehicle’s body. Therefore the design of powertrain mounting significantly because this would require changes to the front systems is governed by specific vehicle requirements. sub-frame and other components. Engine mounting for a heavy duty vehicle has been studied A finite element model and an optimization scheme are by Iwahara and Sakai [1]. The best mounting layout to discussed for the response of this passenger vehicle produce vibration isolation and to support engine dynamic powertrain / powertrain mounting system. The vibration of torque was investigated by using an eigen-analysis and FEM the standard powertrain configuration is analyzed through its simulation. The rigid body modes, frequency response and rigid body mode shapes and an optimization scheme is transient response were obtained and several mounting suggested. The scheme aims to improve the performance of layouts were discussed. It was found that it is possible to the mounting system in order to produce better powertrain attain a mounting layout that will give best performance for overall NVH performance. idling engine vibration and shock due to engine torque.

In reference [2], Li-Rong et al. presented a finite element model of a lumped parameter model for a hydraulic damped ® engine mount. The FEM model can be used in ADAMS PART X (MM) Y (MM) Z (MM) simulations to determine mounting effect on vehicle ride. Front mount 1238.6 67.0 487.2 Rear mount 1829.8 282.2 416.4 Sakai et al. discussed mounting system optimization for Transmission mount 1614.7 -442.6 412.6 vibration isolation and interference avoidance in reference TS bushing – engine 1360.1 499.4 942.7 [3]. An optimization method was presented which allowed for the inclusion of engine packaging as part of the design TS bushing – body 1620.2 499.4 892.2 objectives. This inclusion aimed to reduce the issue of increased stiffness at certain vehicle travel conditions. (Good Table 3 – Baseline powertrain mount locations vibration isolation at low engine amplitudes and poor isolation at high engine loads.)

PART X (N/mm) Y N/(mm) Z (N/mm) In the present , a powertrain mounting system for a large passenger automobile is investigated and an Front mount 205 365 469 optimization scheme for better performance is derived. Rear mount 112 456 456 Transmission mount 160 320 480 Powertrain and Powertrain Mount Specification TS bushing – engine 80 16 80 TS bushing – body 250 250 25 The rigid body mode shapes for the powertrain are obtained through an FEM simulation approach [4]. The resulting frequencies and modes shapes can then be analyzed and Table 4 – Baseline powertrain mount stiffnesses changes can be applied to the stiffnesses of the mounts, as well as their locations, in order to obtain better performance. The vehicle utilized for this study is a front wheel drive with a Figure 1 shows a sketch of the sub-frame and powertrain transversely mounted V8 engine. The engine is mounted on mounting system. an isolated sub-frame that is attached to the vehicle’s body via six elastomeric bushings. The engine is attached to the front sub-frame via three mounts. A torque strut is attached to the engine and to the body of the vehicle also through elastomeric bushings. Table 1 summarizes the powertrain mass and inertia properties in the powertrain inertia coordinate system (Table 2).

CENTER OF MASS (MM) X=1544.9; Y=22.2; Z=603.4 Mass (Kg) 360 2 7 J xx (Kg mm ) 1.74 x 10 2 7 J yy (Kg mm ) 2.48 x 10 2 7 J zz (Kg mm ) 3.10 x 10

Table 1 – Powertrain properties Figure 1 – Sub-frame and powertrain system

Mounting System FEM Model POWERTRAIN INERTIA CO-ORDINATE SYSTEM (MM) The powertrain with its mounting system is modeled using Origin X=1544.9; Y=22.2; Z=603.4 MSC NASTRAN finite elements software package. In the X-Axis X=1544.514; Y=23.045; Z=603.751 finite element model, the powertrain mass and inertia Z-Axis X=1544.612; Y=22.398; Z=602.498 properties are included via a concentrated mass element ‘CONM2’ at the grid point for the center of mass co-ordinate. Table 2 – Powertrain location The mounts are modeled using scalar spring elements ‘CELAS2”, and the mounts are connected to the grid point of the center of mass using rigid elements ‘ RBAR’ or ‘RBE’. A Table 3 gives the locations for the powertrain mounts and schematic representation for the FEM of the overall Table 4 shows the mount stiffnesses of the baseline design. powertrain system is shown in Figure 2.

Design Objectives

The modal characteristics of the powertrain system show the six rigid body modes and their corresponding natural frequencies [6]. Identifying these rigid body modes of the powertrain is one of the main factors in determining powertrain suspension design. Automotive powertrain suspension design is a compromise between isolation of powertrain rigid body modes from vehicle rigid body mode targets and packaging constraints. So, it is very important to decouple the rigid body modes and design the mounting system for frequencies that could prevent amplification. Amplification is caused when a powertrain vibration frequency matches a vibration frequency of some other vehicle assembly. The main design criteria for optimizing the mounting system are [7]:

• Decoupled vertical mode with a frequency between 8 to 9 Hz together with at least 90% of the in the vertical direction. These requirements keep the vertical mode below the suspension wheel hop frequency and avoid amplification. • Decoupled pitch mode frequency between 10.5 to 11.5

Hz with 80% of the kinetic energy in the pitch direction. Figure 2 – Powertrain FEM (top view) This requirement is mainly to avoid pitch-fore/aft coupling and to prevent the frequency range from falling

between the cold and warm idle engine speeds, so that Analysis of Baseline Design vibration from any first order engine imbalance is not

amplified. The modal characteristics of the powertrain system of the • The remaining rigid body modes, i.e. fore-aft, lateral, baseline design were obtained using MSC NASTRAN. The roll, and yaw should have around 75% kinetic energy in natural frequencies of the rigid body modes are listed in their primary direction. Table 5 and the distribution of the modal kinetic energy [5] is shown in Table 6. • Roll and yaw frequency should lie between 7 to 15 Hz. This requirement is to ensure that the lower end of the range is above human sensitivity, while the high end of FREQUENCY (HZ) PREDOMINANT MODE TYPE the range is below any body structure flexible body 5.21 Fore / Aft mode. 7.37 Lateral 9.63 Vertical Design Optimization 10.12 Yaw 11.40 Pitch The baseline design does not satisfy all design objectives for 14.60 Roll the powertrain rigid body mode requirements in terms of frequencies and decoupling.

Table 5 –Baseline mode shapes and frequencies In order to satisfy the powertrain rigid body mode requirements the mount stiffnesses and mounting locations are optimized based upon the design criteria. Achieving the Natural %Kinetic energy distribution desired vertical mode frequency requirement can be obtained by changing the stiffness of the powertrain mounts. Frequency Fore/aft Lateral Vertical Roll Pitch Yaw For the baseline model, MSC NASTRAN design optimization analysis ‘SOL 200’ is used to find the optimum value of the stiffness in the Z-direction for the three powertrain mounts on 5.21 92.3% <1% <1% <1% 5.8% 1.5% the sub-frame. Table 7 shows the optimized values of the Z- 7.37 <1% 73.6% 5.0% 17.8% 1.4% 1.7% direction stiffness for the three mounts. The natural 9.63 2.4% 6.0% 60.3% 1.9% 6.4% 22.9% frequencies and the distribution of the modal kinetic energy for the modified mount configuration are shown in Table 8. 10.12 1.4% 4.8% 24.8% <1% <1% 68.4% 11.40 3.3% 4.0% 7.2% 3.3% 76.3% 6.0% PART X (N/MM) Y N/(MM) Z (N/MM) 14.60 <1% 11.5% 2.5% 76.8% 9.5% <1% Front mount 205 365 496 Rear mount 112 456 327

Transmission mount 160 320 369 Table 6 – Baseline modal kinetic energy distribution Table 7 – Modified mount stiffnesses

Fore/Aft Lateral Vertical Yaw

Pitch Roll %Kinetic energy distribution 100% RBM Fore/aft Lateral Vertical Roll Pitch Yaw Frequency 95%

5.16 91.5% <1% <1% <1% 6.8% 1.4% 90%

6.97 <1% 68.2% 5.4% 25.2% <1% <1% 85% 9.00 <1% 5.8% 88.0% <1% 4.0% 1.6% 80% 9.93 4.4% <1% <1% 2.0% 9.4% 84.1% 11.00 3.3% 7.6% 2.8% 3.6% 70.2% 12.5% 75%

70%

13.45 <1% 18.4% 3.5% 69.0% 9.2% <1% Kinetic Energy (%)

65% Table 8 – Distribution of Modal Kinetic Energy 60% 220 240 250 260 270 282.15 290 300 310 Design Sensitivity of Mount Location Rear Mount Y co-ordinate

In the modified design, the lateral, roll, and pitch modes do not satisfy the decoupling requirement. Therefore, an investigation of the effect of changes in mount location is carried out. Design sensitivity analysis using MSC Figure 4 - Kinetic energy distribution of the RBM for NASTRAN ‘SOL 200’ is used for this location optimization varying rear mount position [8]. The objective of the analysis is to improve rigid body mode decoupling in the system without affecting the vertical and pitch frequencies too much. The design parameters for Design sensitivity analysis for the location of the front mount this design sensitivity analysis are: moving front mount within shows that the vertical and pitch frequencies remain almost ± 100 mm along Y-axis, moving rear mount within ± 100 mm unchanged from the modified system. along Y-axis and moving transmission mount within ± 100 mm along X-axis. Figure 3 shows the kinetic energy distribution of the rigid body modes (RBM) for varying front mount position. It is seen that decoupling occurs as the front mount moves away Fore/Aft Lat eral Ver t i c al from the center of mass. The decoupling of the pitch mode Yaw Pit ch Roll increases when the mount moves from 67.00 mm to 162.50 mm after which decoupling decreases (see Figure 3). Thus, 100% moving the front mount to the new y-coordinate 162.50 mm is desirable in order to decouple the pitch mode in the pitch 95% direction.

90% Design sensitivity analysis for the location of the rear mount

85% shows that the vertical and pitch frequencies remain almost unchanged from the modified system. 80% Figure 4 shows the kinetic energy distribution of the RBM for 75% varying rear mount position. Although decoupling can be achieved through rear mount displacement, due to Kinetic Energy % 70% packaging requirements for this case, it is desirable to retain the original mount location for the rear mount. 65%

60% Figure 5 shows the kinetic energy distribution of the RBM for 40 67 80 100 120 140 160 180 200 varying transmission mount position.

Front Mount Y co-ordinate Design sensitivity analysis for the location of the transmission mount shows that mode decoupling of all the modes are accompanied by a change in the vertical and pitch frequencies. So it is not desirable to shift the transmission mount position. Figure 3 – Kinetic energy distribution of the RBM for varying front mount position Fore/Aft Lateral Vertical Yaw Pitch Roll PART X (MM) Y (MM) Z (MM)

100% Front mount 1238.6 162.5 487.2 Rear mount 1829.8 282.2 416.4 95% Transmission mount 1614.7 -442.6 412.6 TS bushing – engine 1360.1 499.4 942.7 90% TS bushing – body 1620.2 499.4 892.2 85%

80% Table 10 – Powertrain mount locations

75% Conclusions 70% Kinetic Energy % 65% An improved powertrain mounting system configuration satisfying the rigid body frequencies and decoupling 60% requirements can be achieved through a two-step process. The two steps are: (i) optimization of the mount stiffnesses in 55% the vertical direction and (ii) change in mount location based 1560 1580 1600 1620 1640 1660 1680 1700 1720 on design sensitivity. It should be noted that packaging Transmission Mount X co-ordinate constraints are very critical in the determination of the final mount locations, while changes in mount stiffnesses must be verified for static and durability requirements.

References

Figure 5 - Kinetic energy distribution of the RBM for [1] M. Iwahara and T. Sakai, The Optimum Layout of varying transmission mount position Engine Mounting by Dynamic Analysis, SAE Technical Papers, 1999.

[2] W. Li-Rong et al., Characteristics of Hydraulically

Damped Rubber Mount of Car Engine, SAE Final Design Technical Papers, 2001.

[3] T. Sakai, M. Iwahara, Y. Shirai, and I. Hagiwara, Considering the packaging constraints and the design Optimum Engine Mounting Layout by Genetic sensitivity results, the final design is obtained by modifying Algorithm, SAE Technical Papers, 2001. the location of the front mount in the Y-direction. This should [4] J. Bretl, Optimization of Engine Mounting Systems improve the decoupling of the mount rigid body modes as to Minimize Vibration, SAE Technical Papers, 1993. shown in Table 9. The new design satisfies the design [5] M. Wamsler and T. Rose, Advanced Mode Shape criteria stated above for improved mounting system. Table Identification via Modal Kinetic Energy Plots 10 shows the final locations for the mounting system. Assisted by Numerous Printed Outputs, MSC

Americas User's Conference Proceedings, 1998.

[6] T. Sakai, Y. Takano, and M. Iwahara, The

Optimum Design of Engine Mounting, SAE %Kinetic energy distribution Technical Papers, 1998. RBM Fore/aft Lateral Vertical Roll Pitch Yaw [7] L. Piancastelli, P. Barnard, and N. Powell, ADAMS Custom Interface for Powertrain Mounting Frequency System Design, ADAMS User Conference 5.21 92.3% <1% <1% <1% 6.9% <1% Proceedings, 2000. [8] G. J. Moore, MSC/NASTRAN Design Sensitivity 7.26 <1% 76.9% <1% 21.0% <1% 1.1% and Optimization, vol. 67: The Macneal - 9.03 1.1% <1% 92.9% <1% 5.1% <1% Schwendler Corporation, 1992. 10.02 1.9% <1% 1.1% 1.4% 9.6% 85.6%

11.23 4.3% <1% 4.9% <1% 78.2% 12.8% 13.36 <1% 21.5% <1% 78.0% <1% <1%

Table 9 – Distribution of modal kinetic energy